A radial piston engine of this type, as is known for example in French Published Patent Application 2,296,778, and therein only a partial compensation of the hydraulic forces acting on the guide body of the piston is possible in as much as the pressure-medium acting on the guide body from outside to inside always acts in the same direction on the guide body, whereas the pressure acting in the opposite direction from inside to outside, by which the guide body is kept in place against the bearing shell, follows in its alignment the swivelling motion of the guide body, that is to say continually changes the alignment, so that the counteracting compressive forces over the swivelling range cannot be compensated.
This is explained in greater detail with reference to FIG. 1, which diagrammatically shows the known design according to the abovementioned French Published Patent Application. The pressure p.sub.B of the pressure medium, fed in for example via passages in the cylinder cover, which prevails in the pressure space 1 above the guide body 2 over the diameter de of this pressure space exerts a force F.sub.H which remains constant during the swivelling motion of the guide body 2, which engages in a hollow piston 3. In the swivelling position of the piston and guide body represented in FIG. 1, there acts on the underside of the guide body 2, which is provided with clearances for the pressure medium to pass through, the same pressure p.sub.B with the resulting force F.sub.K, which keeps the guide body 2 in place with its spherical-annular bearing face 4 against a spherical-annular bearing face 5 of the housing or cylinder cover. Resolving this resultant, a force F.sub.KY opposes the bearing relieving force F.sub.H acting on the upper side. The component F.sub.KY of the force pressing the guide body against the bearing acts transversely with respect to the longitudinal axis of the guide body 2 and consequently acts transversely on the piston 3.
The equilibrium of the forces in this swivelling position gives a dependence between the permissible degree of bearing relief m and the geometry of the engine as EQU m.sub.por =F.sub.H /F.sub.K =cos .alpha.-sin .alpha.tg .phi..
For example, for a swivelling angle .alpha. of 10.degree. and .phi.=35.degree., a permissible degree of relief of m.sub.por =0.863 is obtained, without taking frictional forces into account.
If at .alpha.=0 the working piston is not swivelled, the guide body could theoretically be relieved completely with a hydraulic counterforce F.sub.H, which is equal to F.sub.K. In this case, the excess of the forces is 1-0.863=0.137, that is to say virtually 14%, producing an adverse effect on the contact pressure on the spherical-annular face between the guide body and the housing, entailing a corresponding frictional moment. An increased frictional moment on the spherical bearing of the guide body causes the shoe of the piston, not shown in FIG. 1, to lift off from the circumference of the eccentric on one side, producing increased frictional and leakage losses in this area, because pressure medium is passed via restricting bores onto the underside of the piston shoe for relief of the hollow piston.
If the frictional forces N. .mu. occurring are also taken into account, the permissible degree of relief is less by a few percent. In order to be able to keep the oscillating guide body reliably in place against the ball seat, a few percent therefore have to be added to this with regard to dimensional and geometrical errors of the ball, so that with the engine data (.alpha., .phi.) specified above an effective degree of relief of about 70 to 75% is obtained.